Fluid drives, which are sometimes called fluid couplings or hydraulic couplings, are commonly used as clutches of sorts to transmit power smoothly between a power source driving an impeller, and a mechanism to be driven, which is connected to a runner. In most of these applications, such as automotive drive trains, the impeller and runner are fluidically coupled in a chamber that is substantially filled with oil at all times, acceleration or deceleration of the driven mechanism being accomplished by the acceleration or deceleration of the engine driving the impeller. These are relatively low power applications, for which the impellers and runners are frequently made of sheet metal stampings or are cast.
All fluid drive impellers and runners have vanes which generally extend radially, defining pockets between them. The cavities formed by the pockets of the impeller and of the runner and the gap between them contain oil. This oil tends to be thrown radially outwardly by centrifugal force, more so by the impeller than by the runner, because the impeller rotates faster. This produces a circulation pattern within and between the pockets of the impeller and runner, where the oil is thrown by the vanes of the rotating impeller against the vanes of the runner, causing the runner to rotate. The leading surface of each vane of the impeller is a high pressure surface, and the trailing surface of each impeller vane is a low pressure surface. As the oil flows radially outwardly through an impeller pocket, it tends to separate from the low pressure surface and to flow with higher velocity along the high pressure surface, forming a jet, one jet per pocket, so that the oil enters the runner as a series of high velocity oil jets. This is in contrast to an ideal uniform flow of oil exiting the impeller and entering the runner.
Two references that cover fluid drives from both the theoretical and practical points of view are Fottinger-Kupplungen und Fottinger-Getriebe, Ernst Kickbusch, Springer-Verlag, Berlin, 1963, and Stromungskupplungen und Stromungswandler, Maurizio Wolf, Springer-Verlag, Berlin, 1962. The effect of fluid performance of planer vanes as a function of the angle relative to the face is discussed in both of these references. They indicate that an angle of approximately 30 degrees to the face maximizes the power achieved from a given impeller diameter.
Conventional fluid drive impellers and runners for high power industrial applications have vanes that are purely radial so that they can be used for either clockwise or counter-clockwise rotational applications, Further, the conventional process of making pockets with over-center milling cutters requires that the sides of the vanes be purely planar.
Generally, the number of vanes, hence the number of pockets, of the impeller is different from the number of vanes on the runner. This is done to spread the pulsing effect of the jets over different vanes, reducing the magnitude of the alternating stresses induced in the vanes. It has generally been considered to be desirable to position the impeller and runner so that the gap between them is as small as possible in order to increase the efficiency of the drive. At the same time, the magnitude of the jet pulsing effect increases with a narrower gap. With a sufficiently narrow gap, this leads to fatigue of the vanes, often leading to their failure. In order to reinforce the vanes, reinforcing rings have been bolted, welded, brazed, or cast intermediate the radial reach of the vanes; the vanes have been made uniformly thicker; and/or the impellers have been made of material with improved mechanical properties. This has been possible in the low power applications, such as automotive fluid drive applications, in which it is possible to use sheet metal or cast impellers and runners. It has also been possible in the industrial fluid drives that are not subjected to severe duty applications.
It should be noted that in some applications, the "reinforcing ring" has a second and perhaps more important function of controlling the oil flow pattern. In the fluid dynamic literature, this ring is called the "core".
The system of this invention has to do with industrial fluid drives transmitting power of a different order of magnitude from that of automotive and other relatively low pressure, low torque applications, usually under severe duty applications. The fluid drives to which this invention relates generally transmit more than 4,000 and up to 15,000 horsepower or higher, per impeller/runner pair at a nominal input speed of 3600 rpm, and generally are subjected to severe duty applications. As used here, "severe duty" implies high torque transmission with high differential speed, or high slip speed, between the impeller rotative speed and the runner rotative speed in continuous operation for years. The problems associated with high powered fluid drives of the type to which this invention is addressed are quite different from those of the low power type and from those which do not experience high differential speed between the impeller and runner.
As indicated just above, the fluid drives to which this invention relates, are designed to provide a variable output shaft speed from a constant input speed. For example, a turbine rotating at a constant speed of 3600 rpm drives a fluid drive which drives a boiler feed water pump at a rotative speed in the range between approximately 3500 rpm and 800 rpm. In this example, if the pump absorbs 14,600 horsepower with the pump speed and fluid drive output shaft speed at 3,500 rpm., then the turbine will supply approximately 15,000 horsepower to the input shaft and impeller of the fluid drive. When the boiler feed pump and fluid drive output shaft rotate at 800 rpm, the pump absorbs perhaps 200 horsepower, with an input power from the turbine of 900 hp. The variable speed of the driven machine is accomplished by varying the amount of oil in the cavity in and around the impeller and runner, conventionally by means of a scoop tube. Parasite losses related to bearings and the scoop tube are not included in the above power figures, which are merely illustrative, and form no part of the invention.
The impeller and runner of a high powered fluid drive to which this invention is addressed are machined from billets or blanks of alloy steel forgings, and this applies to both the conventional designs and the improved designs of this invention. Conventionally, the free edges of the vanes of the impeller are notched to receive a reinforcing ring that engages the vanes at the free edges, with long bolts retaining the reinforcing ring into the impeller vanes. However, the bolts tend to break or loosen, and their use has proved to be hazardous to the long term operation of the fluid drive. In all non-severe duty applications, the vanes are generally not provided with reinforcing rings. Prior to this invention, the use of the reinforcing rings that are bolted in place has been considered the best design for minimizing the probability of breakage of the vanes under severe-duty applications, that is, when high differential speed conditions exist.
The metallurgical properties of the alloy steel forgings and the severe duty of these impellers and runners do not permit welding, brazing, or similar attachment of reinforcing rings to the vanes.
In all fluid drives, heat is generated in the oil due to the inefficiency of the process, being related to the torque transmitted times the differential speed. In the case of these high power fluid drives, the power loss, or the heat generated in the oil, can amount to 200 to 4000 horsepower per circuit, that is, per one impeller and its mating runner. Accordingly, a continuous flow of oil through the circuit is used to carry away the heat, otherwise it would overheat so severely as to be unserviceable.
Conventionally, there are several methods for delivering oil into the circuit, that is, into the impeller and runner cavity: One method is for the oil to enter either through hole(s) in the input shaft and/or hole(s) in the output shaft and then outwardly into the gap between the runner and impeller. Another is for the oil to enter a collection ring attached to and rotating with the impeller, commonly called an impeller oil pump, and then through holes in the shroud. Nothing of this improvement affects the method of delivering oil into the circuit.
Conventionally, the oil leaves the circuit through the gap between the outer surfaces, or shrouds, of the impeller and of the runner, and/or through holes in the shroud of the impeller or runner at some distance from the free edges of the impeller or runner vanes, respectively.
However, certain aspects of this invention address the flow path and the sealing of the oil as it exits the circuit. Conventionally, the surface which forms the shroud on the outer perimeter of the impeller pockets terminates at the plane transverse of the axis of rotation of the impeller which contains the free edges of the impeller vanes. Conventionally, for the runner shroud, there are two designs. In one design, the shroud on the outer perimeter of the runner pockets terminates similarly to that of the conventional impeller, that is, the runner shroud terminates at the plane transverse of the axis of rotation of the runner and which is defined by the free edges of the runner vanes. In the second design, the shroud terminates in a lip in the form of a Vee, with the sharp point of the Vee protruding approximately 1/8" beyond the transverse plane containing the free edges of the runner vanes.
Conventionally then, in the first design, the gap between the impeller and runner shrouds is the same as the gap between the free edges of the impeller and the runner vanes. In the second design, there is a Vee shaped lip on the runner, so that the gap between the shrouds is slightly (e.g. 1/8") smaller than the gap between the free edges of the vanes.
In addition to acting as a structural support for the vanes, the shrouds serve as a seal on the flow of the high velocity, high energy oil, as it exits through the gap between the impeller and the runner shrouds. Clearly, the smaller the gap, the better the seal, and the more efficient the circuit is. This is the justification for designing this sealing gap so that it is no larger than necessary (a) to prevent contact of the impeller and runner due to (1) vibration, (2) axial movement in the thrust bearings, and/or (3) thermal growth, and (b) to pass only the oil flow desired through this exit flow path.
On the other hand, as indicated above, the size of the gap between the free edges of the vanes influences the magnitude of the oil jet pulsing effect as the oil passes from the impeller to the runner and again, from the runner to the impeller.
Conventionally, for the industrial fluid drives, the pockets have been formed by several methods. One is by casting; another is by welding, brazing or similarly attaching the vanes to the shroud or pocket casing, and another is by milling out with a circular, multi-toothed milling cutter which can reach over the center of the cutter.
In the casting method, a pocket of any shape can be made, and it can contain a reinforcing ring, cast in, or, depending upon the material, one can be brazed in or welded in. The problem with cast materials is that they do not have the high strength mechanical properties suitable for the severe duty applications to which this invention is addressed.
In the second conventional method, which uses welding, brazing, or the like, a pocket of almost any shape can be made with vanes of almost any shape or orientation, with or without a reinforcing ring attached. However, welded joints, brazed joints or similarly made joints are not adequate, because experience indicates that they will crack and subsequently fail under the severe duty applications addressed in this invention. In the third conventional method, high strength alloy steel blanks are used to obtain suitable mechanical properties, into which the pockets are milled in slices, each slice being the thickness of the milling cutter, on the order of 3/8 inch (9.525 mm) to 1/2 inch (12.700 mm), two to four slices per pocket, overlapping at the radially inward end of the pocket, and fanning out at the radially outward end of the pocket. This causes the sides of the vanes to be purely planer and conventionally of uniform thickness in both the radial direction and in the axial direction. Because the milling cutter is circular in shape, the bottom of the pocket made by this method is substantially semi-circular in shape. Further, the radius between the pocket bottom and the vane is the same everywhere. Clearly, it is not possible to have an integral reinforcing ring in an impeller made by this method. While this method leaves a compressive layer in the surface as a result of the machining process, which is good, it also leaves tool marks or scratches which must be removed by hand, as these tool marks are very significant "stress risers", particularly in the fillets, from which many cracks have emanated, causing failure of the vanes. When a tool mark occurs, it typically occurs along the entire path that the tool takes along the surface of a vane, the tool forming a generally semi-circlular shaped pocket. When such a tool mark occurs in a fillet, it usually occurs along the entire length of the fillet between the side of a vane and the pocket bottom. These tool marks must be removed, and are usually removed using hand tools such as pencil grinders, leading to under-cuts, or thin spots, in the vanes at the intersection of the vanes and the pocket bottoms. These thinner spots in the vanes are themselves stress risers, which also have contributed to vane failures.
One of the objects of this invention is to provide an impeller for a fluid drive system adapted to handle 4,000 hp up to 15,000 hp or more, and to deliver variable speeds over a wide differential speed range, including those described as severe duty: a differential speed of 100 rpm to 2800 rpm, for an input speed of 3600 rpm, and in which the vanes are reinforced in such a way as substantially to eliminate the danger of breakage of the reinforcement.
Another object is to provide such reinforcement that causes little interference with the circulation pattern of the oil.
Another object is to provide an impeller and a runner in which the magnitude of the jet pulsing effect is reduced, while at the same time, providing for maximum sealing of the high velocity, high energy oil as it exits the impeller.
Another object is to provide such a fluid drive in which the impeller vanes, and if desired, the runner vanes are formed to optimize their strength.
Still another object is to provide such a fluid drive in which the impeller vanes, and if desired, the runner vanes are formed to optimize the transition of the oil flow into and from the pockets, hence smoother acceleration or deceleration of the fluid and corresponding transfer of power to the oil and from the oil than has been provided heretofore in fluid drives of the type to which this invention pertains.
Other objects will become apparent to those skilled in the art in the light of the following description and accompanying drawings.